Variable delivery pump or motor



v /f/L/ Dec. 3, 1935. E. K. BENI-:BEK 2,023,216

4 l VARIABLE DELIVERY PUMP OR MOTOR Filed Jan. 29, 1932 5 sheets-sheet 1 Dec. 3, 1935. E. K. BENEDEK VARIABLE DELIVERY PUMP OR MOTOR 9/w @y ,W /M fam //////////f//////////// w .W WJ m f w W ,MMM /w fa d Y m w1 Mm C fr0/ w M K 7% wf.. ik .n 97 W ao k W mm @plm attorneys.

Dec. 3, 1935. E. K. BENEDEK 2,023,216

VARIABLE DELIVERY PUMP 0R MOTOR Fiied Jan. 29, 1932 5 sheets-sheet 3` ZP" 60 i" .A4 w

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Dec. 3, 1935. E. K. BENEDEK VARIABLE DELIVERY PUMP MOTOR Filed Jan. 29, 1932 5 Sheets-Sheet: 4

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Dec. 3, 1935. E. K. BENEDEK 2,023,216

VRIABLE DELIVERY PUMP OR MOTR Filed Jan. 29, 1932 5 Sheets-Sheet 5 Patented Dec. 3, 1935 UNITED' STATES- VARIABLE DELIVERY PUMP OR MOTOR Elek K. Benedek, Mount Gilead, Ohio, assigner to The Hydraulic Press Manufacturing Company, Mount Gilead, Ohio Application January 29, 1932, Serial No. 589,746

5 Claims.

This invention relates to rotary radial pumps or motors in which rotors revoluble about parallel axes are positively connected through piston and cylinder assemblies to cause drive of one rotor by the other through such connections.

Pumps of the type referred to include primary and secondary rotors rotatable about parallel axes and shiftable relatively to vary the relative eccentricity of one to the other for varying the output of the pump. Radial piston and cylinder assemblies are the connections between the rotors to engage the rotors for synchronous rotation. Loose connections even though slight, are sufficient to upset perfect dynamic balance of the rotating parts and cause dynamic vibrations,y noise, unnecessary wear and breakage.

It is the purpose of my invention to provide a balanced pump operating at high speed and high pressure, in which a primary and a sec'- ondary rotor are connected and balanced by novel means, to provide for maximum load carrying capacity through well lubricated bearing surfaces, and one which efectively eliminates fric-- tion due to misalignment, and all loose connections Which produce impact, or phase differences under conditions of operation. Another object of the invention is to provide a novel integral piston and cross head unit in which the cross head is equipped with adjustable means to permit, in assembly, self alignments of the units without undue operative play; and in which means is provided for directing the hydraulic piston and reaction load. forces to a single focal point.

Another object of the invention is to provide means in combination with the cross heads in which the load transmitting surfaces are so formed as to provide high pressure wedge shape oil films at all times between said means and reaction surfaces.

Another object of the invention is to provide novel means to balance the synchronously rotating primary and secondary rotors, especially when the pump is set at the maximum stroke and the rotating masses are in eccentric relation,'thereby to improve the performance and life of the pump.

Other objects will in part be obvious and in part be pointed out hereinafter.

To the attainment of the aforesaid objects and ends, the invention still further resides in the novel details of construction, the combination and arrangement of parts, all of which will be rst fully described in the following detailed description, then be particularly pointed out in the ap- (Cl. 10S-161) Figure 1 lis a vertical cross section of the pumpI taken through the main meridian plane of the pistons, part being shown in elevation and part in section.

Figure 2 is a horizontal section taken through the axis of the pump, parts being shown in section and in elevation.

Figure 3 is a detail fragmentary vertical cross section taken through the axis of one cylinder and piston assembly and illustrates the self-aligning quality of the cross head when the reaction mem-- ber is out of alignment.

Figure 4 is a detail fragmentary vertical longitudinal section taken through the axis of one cylinder and piston assembly and shows a slightly modified form of cross head in which the adjustable surfaces of the cross head constitute a portion oi a Vspherical surface with the center line on the axis of the cylinder barrel and serve to direct hydraulic reaction forces to a focal point in the centerline of the pistons.` Y Figures 5, 6, and '7 illustrate the outer element of the two-piece cross head shown in Figure 3 in central longitudinal section, in plan and in central transverse section respectively.

Figures 8, 9, and 10 illustrate the adjustable element of the cross head shown in Figure 4 in central longitudinalr section, in plan and in central transverse section respectively.

Figure 11 is a diagrammaticdetail view illustrating a piston unit of a Well known conventional form of Apump in which the load is carried on a .Wrist pin carried by slippers and which are subject to wear ras diagrammatically illustrated in Figure 11a.

Figure 12 is a view similar to FigureI 11 showing another well known conventional form of A pistonv unit wherein the load is carried on a miniaturepush pin. t

Figure 13 is a comparative view similar to Figures l1 'and 12 and illustrates my improved piston unit in a manner permitting a comparison of the load transmitting means with the conventional units mentioned in regard to friction and construction.

Figure 14 is a diagrammatic cross sectional view illustrating the non-synchronous action which'can take place whenthe primary and secondary rotors are connected by flexible devices such as are illustrated in Figure 12.

Figure 15 is a comparative view similar to Figure 14 and shows the effect of employing rigid piston units of the type shown in Figures 3 and 4.

Figure 16 is a diagrammatic cross section illustrating the relative positions of a piston unit and rotors in vseven successive positions during one revolution of my improved pump structure.

Figure 17 is a cross section illustrating the relative positions-of primary rotor and a radial series of seven pistons.

Figure 18 is a diagrammatic view of the secondary rotor and illustrates relative positions of secondary rotor and a radial series of seven pistons. Y

Figure 19 is a composition of Figures 17 and 18 and illustrates the relative positionof the piston units and both the primary and secondary rotors. l

Figure 20 is a detail vertical cross section taken on the line 28-20 on Figure 2.

Figure 21 is a detail perspective view of the control rods and assembled bearing rings of the secondary rotor.

Figure 22 is a detail sectionalview of a cross head and discloses means for limiting the adjustment of the associated parts.

In a practical embodiment of the invention I provide a pump casing comprising a cylindrical ring 5 having integral supporting feet 6, and upper and lower parallel guide pads 1.. The cylindrical casing is4 also provided with diametrical control hubs in which shifter rod bores 8 are formed, and provided with drain back ducts 9 which communicate the interior of the casing to enable fluid to leak past the control rods.

The cylindrical cover is completed by end casings I ll and I I, which are centered on shoulders I2 in regard to casing ring 5 and which are flanged and secured as at I3 to the respective ends of the casing ring. Each end cover is also provided with an annular recess I4 adapted to receive a sealing and bearing ring I5. As will be observed by reference to Figure 2 the rings I5 are apertured to receive and enable them to slide axially on screws I6 secured in the end covers. In the space between each ring I5 and the adjacentend cover recess wall, spring elements I1 are inter-,- posed and serve to hold the rings toward the rotors to close the pump chambers formed within the casing ring 5 and end covers ID and I I.

The end cover I0 is provided with an axial bore I8 in which the enlarged head I9 of the cylindrical pintle 20 is rigidly secured and positioned by means of a key 2I. The end cover I0 is provided with diametrically opposed intake and exhaust ports 22 and 23 respectively which lead into the bore I8 and register with ports 24 and 25 formed in the enlarged head of the pintle. See Figures 1, 2, and 20;

The pintle port 24 communicates with a pair of longitudinal ducts 2,6 provided in the pintle, and the port 25 communicates with a similar pair of ducts 21 formed in the pintle. The pairs of ducts 26 and 21 extend only part way through the pintle and have their outer ends plugged as at 2 8 so that communication with said ducts through the pintle head may be had only through the ports 22-'25. It is understood,.,that each of the pairs of ducts 26 and 21 are'intake and exhaust ducts respectively, according to the adjustment of the pump during the particular interval.v

At their ends opposite the plugs 28 the upper and lower pairs of ducts 26 and 21- communicate with upper and lower pintle'ports 29 and 30 respectively, which afford comm: nication between move in varying the output of the pump always 10 coincides with the 'center line of the bridge pieces.

The end cover II is axially bored and is counterbored to receive a bearing sleeve 4D. The bearing sleeve 40 forms a plain film bearing for the head 4I of a drive shaft 42 which extends from l5 the casing through a packing gland 43 for attachment with suitable driving means not shown. The gland bearing 44 is flange-secured to the end bell II.

A cylinder barrel or primary rotor 45 is rotat- 2o ably mounted in the pump casing, having its principal bearing, a plain oil film bearing, on the main cylindrical body portion 20 of the pintle. The primary rotor is engaged at its ends by the sealing and bearing rings I5 which are pressed 25 slightly against the end faces thereof .by the spring elements I1.

The primary rotor 45 is provided, in its central transverse plane, with avplurality of radial cylinde'i bores 46 spaced equidistantly and open 3f) radially. At their inner ends the cylinder bores 46 communicate, through ports 41, with the pintle and the ports 29 and 30. Thus, as the primary rotor is rotated about the pintle each cylinder bore 46 thereof will communicate alternately with the pintle cut-outs 29 and 30 and the upper and lower pairs of ducts 26 and 21 respectively.

At its closed end the primary rotor is provided with a cross groove, and the drive shaft head 4I is provided with a similar cross groove lying in 40 a plane normal to the plane of the rotor groove, and these grooves are adapted to receive correspondingly positioned coupling ribs extending from the opposite faces of the coupling head 5I which is interposed between the cylinder barrel 45 and the shaft head 4I and which is disclosed in detail in my copending application No. 564,404.

A pair of shifter rings 54 are mounted within the pump casing and are rigidly connected and held in the spaced relation shown in Figure 2 by 50 yokes 55 one thereof being provided at each side of the secondary rotor, and each including spacer shoulders 56. The yokes 55 are secured by screws 51 or otherwise to the rings 54 and to these yokes the shifter rods 58 reciprocable in the guides 8 are 55 secured in any approved manner as indicated at 59. Ea'ch ring 54 is provided with a pair of slide guide pads 60 which engage the pump casing pads 1 and control shifting movement of the rings so that their axes and the axis of the secondary rotor having rotary bearing therein always align and lie in the line of the axis of the pintle and center line of the bridge Aportion 3|. The yokes. 55 have their outer surfaces curved to accord in shape with the surfaces of the casing ring 5 which they oppose so as to adapt them to use as limiting stops The numeral 6I designates the surface of the ring or bearing member 54 on which is formed a 70 high pressure lubricant film as disclosed in my copending application for patent filed August 18, 1931, Serial No. 557,888.

Surrounding the primary rotor 45 is an eccentric'or secondary rotor 62, the center of rotation 75 of which is4 adjustable relative to the axis of the primary rotor by means of the shifter rings 54. The secondary rotor 62 surrounds the primary rotor 45. The secondary rotor comprises a ring having a deep circumferential channel 65 comprising successive tangential sections and having radial walls or flanges 64. Within this channel move piston heads 16--11 of the connecting as-` semblies, the stems or bodies 14 of the pistons extending through elongated perforations or openings 68. To adequately assure retention of ample bath of lubricant, a lubricant retaining ring 66 is secured in place by screws 61 or the like to com- -pletely close or seal the channel 65 in a radially outward direction. It will be observed by reference to Figures 1 and 2 that the annular recess 65 communicates with the interior of the secondary rotor 62 through elongated openings 68, one of which is provided for each cylinder bore provided in the cylinder barrel 45.

The provision of the central shoulder enlargement on the secondary rotor 62 provides lateral bearing hubs or flanges 69 which may be equipped with bearing rings'10 in cooperation with the oil film bearings 6i. The secondary rotor also has lateral engagement with the bearing and sealing rings I5 for the purpose of retaining lubricant as hereinbefore described.4 'Ihe hubs are apertured as at 1l so that free passage of lubricating fluid to the oil lm bearings 6I is provided for.

It will be observed by reference to Figures 1 and 2 that the secondary rotor 62 is provided,

between the flanges thereof which definev the secondary rotor 62. Each of the pistons has an integral cross-head which, together with the block 16, constitutes a piston cross head unit.

In the form of the invention'shown in Figures 1, 2,'3, 5, 6, and '1, the outer face 11 of each piston T-head is in the nature of a convex surface that is cylindrical transversely and arcuatel or convex throughout its length, the centers of curvature lying on an axis perpendicular to the axis of the piston, this axis preferably coinciding with the pintle axis. The mating face 18 of the block 16 is shaped in the form of a concave surface in suitable manner so as to engage the cross-head face throughout the whole area. By

' reason of this geometry the reaction forces acting on the piston cross heads are, in every position thereof, directed to a focal point in the axis of the piston of the particular unit and in the intersection with the axis of the pintle or primary rotor.

The intersection of the pintle axis with the main meridiani-of the pistons might be termed as the focal point in the pump. I nd it advantageous to arrange the curved mating surfaces of the cross head elements 15 and 16 so that the common center of curvature of all of the cross heads will coincide with the pintle axis at all relative positions.

In Figures 4, 8, 9, and 10, I have disclosed a modified form of cross head element engaging surfaces in which the mating faces 11a and 18a form a part of a spherical surface having its center of curvature lying in the radial or axis of the respective piston and in the intersection point with the pintle axis. These surfaces concentrate or direct the load forces to a focal point lying in the piston axis as above outlined.

The upper or outer face of each cross head element 16 engages a stationary reaction plate 80 having lateral tenons 8| extending into the tangential grooves 12 provided therefor between the flanges of the secondary rotor and each such plate is provided with a threaded tap to receive the screw 82 by which a lock plate 83 having its ends extended into the radial grooves 13 of said lo ring or eccentricfis secured to position upon said reaction plate. By reason of the extension of the ends of the locking plates into the radial grooves 13, they serve to secure the reaction plates against tangential movement relative to the crosshead 15 elements 15, 16. The tenons 8| serve to hold the reaction plates against radial movement and thus enable them to take up the radial thrusts of the pistons 14.

The casing ring 5 may be provided with a20 drain 84, if desired, so that fluid leaking into the. pump chamber surrounding the shifter rings 54 may be drained from the pump casing.

It will be understood that with the pump parts positioned as illustrated in Figure 1 of the draw- 25 ings, with the shifter ring center coinciding with the axis of the pintle, the primary rotor 45, the secondary rotor or eccentric 62 and connected parts will rotate freely within the casing without effecting any pumping action. In other words, Figure 1 illustrates a neutral position of the pump.

By means of the shifter rods 58 and yokes 55 the shifter rings 54 may be shifted so that the center line of the eccentric 62 will be moved to-35 one side or the other of the center line or axis of the pintle 20. Let us assume that the ring is shifted to the 'right in Figure 1, and that the direction of rotation of the cylinder barrel 45 is clockwise. In this position of the parts all 40 of the pistons above the center line will be moved radially outward and will 'suck iiuid through the intake port 22 and the upper pair of ducts 26, and the pistons below the center line will be moved radially inward and will expel fluid through the lower pair of ducts 21 and the exhaust port 23.

Obviously, when the eccentric 62 is shifted toward the left of the center line of axis of the pintle the cycle will be reversed and the pistons above the center line will expel fluid while those below the center line will draw fluid into the cylinder bores.

It will be understood that whenever the eccentric 62 is adjusted off center and the primary 55V rotor 45 is rotating, the pistons 14 will be reciprocated radially in their receiving bores, the' cross heads 15, 16 oscillating tangentially in the secondary rotor recess 65 because of the off center relation of the secondary rotor. 00 The cross heads shown in Figures 3 and 4 .-.are made as large as possible so that themultidimensional mating reaction surfaces of the component parts thereof are of maximum area.

thereby to minimize the unit pressure per unit 65 area and consequently reduce to a. minimum wear and piston knock.

mating surface of the cross head elements and 16 preferably form part of a cylinder or of a sphere respectively, with their centers of curvature always lying in the axis or radial of the respective piston unit and as close to the pintle axis as operation will permit.

In order to obtain maximum cross -head strength and to properly direct reaction forces, I prefer to make the cross head surfaces 11 convex.' But they may be of any other form which will concentrate the reaction forces in a focal point lying in the centerline of the cylinders.

Whenever the receiving grooveways 65 are formed with error so that a rigid cross head element could not accurately bear against the re- 'action face of the plate 80, my improved solidadjustable cross head will act automatically to compensate for this error and assure the desired full surface contact between cross head and reaction plate.

In Figure 3 I have shown how the novel cross head unit self-aligns under conditions above outlined. The element 16 has only to shift to the right or to the left as indicated in order to permit the reaction faces of 16 and 80 to fully engage. This tangential self-alignment is usually satisfactory for pumps wherein a single row or only a few rows of pistons have been used. However, in a multi-row pump design, where an axial deection of the secondary rotor is unavoidable, a lateral self-alignment is desirable and it is for such large pump design, that a lateral self-alignment is provided for as shown in Figure 4, Wherein a spherical surface form of cross head surface is disclosed. See also Figures 8, 9, and 10.

While I have disclosed simple spherical and cylindrical surfaces, I wish it to be understood that the scope of my invention embraces such other surfaces, other than above said surfaces, which are capable of like self-aligning and load carrying functions.

By reason of their cooperative curved surfaces the cross head elements 15 and 16 in effect constitute a single cross head block and act as one piece except for the relative adjustment for selfaligning purposes.- Therefore, the element 16 follows the harmonic oscillation of theA piston crosshead unit 15 and there is no clearance between the cooperative surfaces 11 and 18 except that which is suflicient to accommodate a thin film of oil which prevents wear of the cooperative surfaces and prevents noise at the pressure change of the piston, especially in the dead centers as the piston change: from suction to delivery stroke or vice versa.

In those critical moments, as it is well known, at high speeds, there are 30 to 50 strokes for each piston per second, or the time for a whole stroke of the piston is about 1/30-1/50 of a second. By further considering the fact that the pre:=ure change on the piston must be only a small fraction of the whole stroke period, it will be evident thatvthe time of such pressure change, by. assuming it to be -1/1000 of the full stroke time, will gure in absolute time measure to 1/29000 of one second. It will be practical to say that the pressure change in the dead centers in high speed, high pressure, rotarypumps of `commercial design, considering that medium sized pumps have to run at a speed more than 860 R. P. M., is approximately 1/30000 (one thirty-thousandth) of one second. It is therefore evident that any slack in the cross head connection must be taken up in an almost infinitesimal time element, that is, with a blow.

In my novel piston construction, the large contact surface in cooperation with a thin oil film eliminates that high frequency noise caused by the take-up of loose cross head connection in the pressure change points of the stroke. It is fur- 5 thermore evident from Figure 3 and Figure 4 that the cooperative surfaces 11, 18 or 11B, 18EL of the cross head, being curved surfaces prevent the tangential escape of oil in the case of the cylindrical surfaces or prevent the escape in all 10 directions in the case of the spherical surfaces. Therefore, in my novel type of pump, provision is made to oil-cushion the piston load in the moment of pressure change and provide a resultant dashpot eiect assuring quiet operation of 15 pump. It is obvious that such dashpot effect cannot occur between plain ordinary surfaces where the oil nlm can be squeezed out at low pressures. i

It will be observed byreference to Figures 1, 5, 20 6, and 7v that the cross heads each include inner plane fac-es 85 opposed to the outer working faces of the secondary rotor tangential grooveways and that the reaction faces of the elements 16 thereof, opposed to the reaction vfaces of the plates 80 25 are provided with a plurality of pairs of inclined plane faces 86 opposedly inclined in the direction of reciprocation of the cross heads. At the meeting edges of each pair of such faces a transverse oil groove 81 is formed as a supply channel 30 for oil.

The shell G6 retains oil thrown into the secondar'y rotor chamber 65 so that the cross heads always reciprocate in an oil bath, the surfaces 86 serving during reciprocation to build up very efficient high pressure oil film bearings. 'Ihus the provision of the retaining shell, the inclined surfaces 86 and the grooves 81 provide for highly efficient lubrication and great load carrying capacity. In application, the slant faced surfaces are interchangeable and the lm bearings afforded thereby serve to eliminate friction and impacts and assist in assuring synchronous rotation of the primary and secondary rotors.

As outlined above the cuwed surface engagement normally assures against displacement of the cross head elements 15 and 16. However. should long usage occasion wear resulting in inefficiency in. this respect, this condition can be guarded against by provision of a ball key 88 50 seating in a hemispherical seat 89 in the element 15 and received in a slightly enlarged recess 96 in element 16 so that self-aligning Amovement of 16 on 1,5 will be permitted but greater displacement of these parts positively limited. See Figure 22.

In Figure 16, A indicates the rotating privmary rotor; B the eccentric or secondary rotor;

and C the piston units which serve as rigid coupling elements connecting the primary and 604 secondary rotors. This diagrammatically illustrated radialgtangential movement kinematic is capable of 'functioning at high speed and high pressure without impact and knocking if the connection of the primary and secondary rotors A and B is rigid. This desirable condition can be satisfied only when the radial path of the piston plunger is at a right angle to the tangential path of the cross head at all times or at all phase angles qb. 'I'his is equivalent toA requiring that 70 the plunger be parallel with the momentary center-line O2 2 of the tangential cross head path. By observing the geometrical relation a-t ther right hand side of Figure 16 as =zero, it will be seen that the center line O1 1 of the primary 75 rotor coincides with the center line of the secondary rotor O2 2 at this instant. From this instant, however, following an anti-clockwise rotation of the assembly as shown by the arrow, the center line of the primary rotor will take the lead, the distance of lead beingl proportional with the trigonometrical sine of the angle qs, that is it is equal to p s in qs, where p is the eccentricity of the secondary rotor.

By following further the rotation of the assembly unit to an angle larger than 90, there is still a relative position of the center lines of the primary and secondary rotors and the center line of the primary rotor still is in a leading position. At 180 however the center lines coincide or merge again. After 180, however, the secondary rotor B or eccentric takes the lead, and this relative condition continues the remainder of the 360 of travel or to 0 degree.

At zero degree, there is a change again, i. e, the

center line of the primary rotor A will again assume the leading position. In' an assembly where the center O2 of rotation of the secondaryl rotor Bi is to the right of the center O1 of rotation of primary rotor A, the rotation being anticlockwise as indicated in Figure 16, the piston center line will have a motion so related tothe center line of the secondary rotor that it Will be in a leading position from to 180 and thereafter in a follow up position from 180 to 360. During all of these positions the two center lines must be absolutely parallel in order to limit the relative positions. At the instances of 07 and 180 the center lines cross each other, and the one will go over to a relatively accelerated motion in regard to the other.

With the above in mind it should be apparent that if at either point of change where the two center lines have to cross each other as above outlined, bne of the rotors is` not in the proper geometrical relation due to exible coordination of the parts of the assembly (see for example Figure 12 and Figure 14, where the primary rotor is in a phase lag of an angle d4 in relation to the driving rotor B) and the lagging element must overcome the lag and as indicated in Figure 14, the piston has to carry over in instantaneous motion in order to come into the theoretical position, after which it may again take up its loose lagging condition this tixne, however, in the opposite dlrection, because of the fact that the primary rotor now assumes a decelerative motion relatively to the secondary rotor. Obviously, this instantaneous position changing occasions great shock and makes itself apparent in the form of noise and breakage.

Shortly summarizing the signicance of Figure 16 it will be seen that during one half of a revolution of the pump assembly the secondary rotor is in accelerated motion relatively tothe primary rotor, Whereas during. the other half ofthe rotation this relation changes, and the secondary rotor assumes a decelerated motion relatively to the primary rotor. The coupling members of the assembly, that is the pistons, have to take up the rotary forces of one of the members if the other is in connection with a driving shaft, driven at uniform speed from a constant speed from a suitable source of power, say for example an electro-motor. It is evident from above consideration that the synchronous rotation of the rotating pump members requires rigid connection between themselves, which positive coordination has been provided for in my present rigid piston unit construction, which is capable of taking up the rotary accelerative forces between cylinder and eccentric eliminates the impacts caused in some structures of the .prior art by free rocking of the crosshead relatively to its plunger, and thereby provides smoother running of the pump 5 inv addition to eliminating torsional oscillation of the coupled rotary members of the pump. The positive attachment afforded by my novel type of self-aligning piston and crosshead units makes it possible to transmit in tune and continuously 1o the synchronous forces from one rotor to the other especially in the change points, where thev forces are zero.

From the distance relation ==psin of the center lines, it follows by the simple rule of mechanics, that the acceleration rc=pw2sin i. e. the acceleration is in direct proportion with the second power of the angular speed w of the pump. The force produced by the impact is further proportional with the mass of the piston unit. The larger the stroke, i. e. the greater the pump is, the greater will be the need for a synchronous coordination of the rotating members during their high speed rotation.

The many advantages afforded by my improved piston and cross head units should be readily appreciable. Flexible connections, such as are shown in Figure 12 cause eccentric loading on the bottom of the hollow piston there shown,4 and the eccentric load results in a tangential component of the hydraulic load, and causes wall friction and wear on cylinder wall. As long as the speed of the pump in proportion to the exibility of the piston is negligible, and the eccentricity and vstroke, of the pump are small, the above-named faults give little trouble at medium pressures. However, at higher pressures, the faults of eccentric piston load occasion serious troubles and in combined high speed and greater stroke designs are intolerable. The application of flexible piston connections used in the prior art arises from the necessity to provide compensation for the deflection of pintle and distortionl of secondary rotor. In order to illustrate this more clearly I have shown in Figures 11, 12, and 13 three different units of properly related proportions and adaptable tc use in pumps of like capacity of output. Figure 13 shows my novel integral rigid piston with selfaligning cross head unit. Figure 11 shows another form of piston selected from a prior art pump of this type. This prior art piston is well known to include two slippers held in cooperative relation on a pin having a diameter d2. The pin naturally has to take all 55 of the hydraulic load of the piston and is of the crank type arrangement. This can not elect positive coupling between primary and secondary rotors, but permits the secondary rotor to float.

In Figure 12, there is shown another prior art piston, in which theaload istransmitted to the cross head by a push pin integral with the cross head and having a diameter cl3.

By a glance at Figures 11, 12, and 13, it will become evident that each piston has a diameters@ D, consequently each is able to suck the same amount of fluid during a given stroke. However, by reference to Figure 13, it will be seen that my load-transmitting capacity has! the full diameter D, whereas the maximum possible pin diameter in Figure 11 is about one half of D, and in addition is exposed to combined stresses. Thus it Will be seen that the section modulus of the pin of Figure 11 is about one-eighth of the section modulus of that in Figure 13 and, therefore, its 7.5

load-carrying capacity is only one-eighth that of my plunger.

In Figure 12, the maximum possible diameter of the pin is only about one-third of that of the pin or section of Figure 13 which makes it only onetwenty-seventh as strong as the section modulus are in a proportion as the third power of the diameters.

Therefore, it should be obvious that myv irnproved unit is able to carry twenty-seven times as much load with the same factor of safety as the unit of Figure 12 and three times as much as the unit of Figure 11.

In order to visualize the balanced condition of my improved pump while in motion, calling it dynamical balance, let us consider the secondary rotor or eccentric as indicated in Figure 15 with one of its pistons located in its groove without definite attachment in tangential direction. By-

considering the fact that the secondary rotor rotates around its concentric bearing at constant speed, it is evident that a piston unit with integral rigid plunger and cross head unit, as shown in Figure 15, will stay in its position of the center line of the rotor, without need of employing any forces other than centrifugal force caused by the rotation and denoted by R2 in Figure 15.

In order that the pistons shall follow the rotation of the primary rotor or cylinder, as shown in Figure' 13, and in order to impart a harmonic oscillation to the piston in the direction of the cross head groove, it is necessary that the primary rotor impart tangential forces T1 and T2 to the piston and thus attain the pure harmonic oscillation in the secondary rotor as illustrated in Figure 15. The tangential motion ofthe piston in the secondary rotor is therefore provided by the primary rotor, and according to the law of equilibrium, the tangential forces T1 and T2 will occasion equal but oppositely directed forces T1, T2 on the shaft of primary rotation as reactions. In a similar manner the centrifugal forces awake reaction forces on the shaft of the secondary rotor. It is evident from above reasoning that the centrifugal forces of the rotary pistons are controlled by the secondary rotorand taken up by the bearing of it, whereas the tangential forces of the pitons are controlled by the shaft of the primary ro or.

In order to illustrate more clearly the aboveA rules, in Figure 19, I have shown the primary rotor A, the center of primary rotation O1, the secondary rotor B, the center of secondary rotation Oz, and seven piston units, all integrall rigid pistons C serving as coupling means between primary and secondary rotors. Considering the secondaryvrotor to have an eccentricity p and a stroke 2p, a careful inspection of this Aillusreference to Figure 17 wherein it will be seenfz;

that the rotary piston forces are greater toward the right of a vertical line intersecting the axis.,

of rotation of the primary rotor.

The effect of rotary piston forces on the balthe co-rotating assembly.

' guide in their respective paths.

ance of the secondary rotor is shown in Figure 18. The rotary piston forces are over-balanced toward the left of a Vertical line intersecting the axis of rotation of the secondary rotor. These drawings alsoshow that the rotary forces are in balance in regard to the horizontal axis N-N in the primary as well as the secondary rotors.

It will be noticed that the over-crowding of the pistons in the one side of the secondary rotor will balance the over-stroking of the pistons in 10 the other side as soon as the piston'plunger and cross head are rigid and integral. As a result of many years of experience and study of this balance problem, I found out that the total increment of the accelerative forces of the eccentrical- 15 ly located pistons acting with the cylinder barrel in tangential direction will be equal to the total -`increments of the accelerative forces of the pis- 30 I will attempt to illustrate further, by comparison how, in my improved pump structure, the combination of the new integral rigid pistons and their perfect alignment producing features assure dynamical balance at all times, thus eliminating 35 torsional vibrations from the pump. vFigure 14 illustrates that a flexible piston assembly such as is shown cannot transmit synchronous rotation of the primary and secondary rotors because of the time lag permitted by the looseness of the piston rod in its hollow piston'. Therefore, the proper geometrical relation, lshown in Figure 16, and wherein the pistons are rst accelerated then simultaneously decelerated in the eccentric cannot be accomplished in such structures as shown in 45 Fig. 14 and the simultaneous values are not equal,

the rotating pistons themselves have no positive As the positionsA of the pistons in such devicesy are exible and determined only by limiting means in their re-n 'spective members, it is obvious that the transmission of balancing dynamical forces between the rotors is not continuous' but intermittent, i. e., shock-like and therefore it causes shocks and vibrations varying in violence according to the degree of looseness and free rocking of the rotating pistons. In the case of use of loose and indefinite con-V nections of rotating systems considerable wear on the pistons is produced by the dynamical forces which cannot be uniformly distributed on all of the pistons and those which are looser than others are not able to transmit their own share and consequently subject the remaining pistons to intolerable stresses. In my many years experience with such units I have witnessed muchbreakage attributable to these inefficiencies. Even with piston units as shown in Figure 1l wherein the pin is larger and has less play than the push pin shown in Figure 12, the enormous rotary forces, or rather the forces of the harmonic oscillation of the cross heads, cause, in a very short time, oblique wear in two right angled directions as shown in Figure 11a, and thus soon `produce loose wrist pins and indefinite rotc connection. In u aoaaaie pumps incorporating such, pistons, dynamic balance can not be effected between the rotors due to the fact that the plunger rocks in regard to the cross head shoe. The seriousness of the existence of unbalanced' forces cannot be over emphasized, as they increase in magnitude at `a quadratic rate with the speed and their effect is cumulative and destructive.

Another advantage of my improved integral rigid piston unit structure resides in the greater suction effort eiliciency occasioned by its use. The suction effort of my improved piston unit is muchvmore steady and positive than in the case of resilient connections of the type shown in Figure 12. In addition the integral pistons have a flywheel effect and provide for a more steady and t uniform rotation of the system, and at Athe same time improve the dynamical balance of the whole system.

In addition also to this flywheel effect and the provision of positive connection between the primary and secondary rotors the great weight of the piston and cross head units enable them to operate during the suction or outward strokes by centrifugal force alone acting on their masses thus permitting them to remain always in cushioned contact with the reaction face of the secondary rotor and avoiding the necessity of changing seats on inner and outer surfaces of the secondary rotor and thenoise and shocks resulting from such seat changing.

Still another advantage of the steady dynamical balance for which I have provided resides in having the central valve or pintle forming the bearing for the primary rotor take up the balanced forces steadily instead of intermittently, thus insuring a steady and continuous oil lm bearing.

From the foregoing description taken in connection with the accompanying drawings it is thought that the novel details of construction, the manner of use and the advantages of my invention will be readily apparent to those skilled in the art to which it relates.

I claim:

1. In a pump or motor, a primary rotor having a cylinder, a secondary rotor having a guideway, a piston reciprocable in said cylinder, and a two-part cross head located within said guideway and having one part rigidly attached to and -movable with the piston, said cross head parts having mating cylindrical surfaces whose axis is parallel to the axis of rotation of one of said rotors.

2. In a pump or motor, primary and secondary rotors rotatable about separated axes, and radial piston and cylinder assemblies connecting the rotors and including cross heads, the pistons being caused to reciprocate in said cylinders by reason of the movement of the rotors about their separated axes, one of said rotors being tan- 5 gentlally grooved to slidably receive said cross heads, and each said cross head being formed of two elements, one element being rigidly attached to its piston, said elements having lcorrespondingly curved relatively-reciprocable cylin- 1o drical mating surfaces whose axis is parallel to the axis of rotation of one of said rotors.

3. In a pump or motor, primary and secondary rotors rotatable about separated axes, and radial piston and cylinder assemblies connecting the 15 rotors and including cross heads, the pistons being caused to reciprocate insaid cylinders by reason of the movement of the rotors about their separated axes, one of said rotors being-tangentially grooved to slidably receive said cross heads, 2o and each said cross head being formed of two elements, one element being rigidly attached .to its piston, said elements having correspondingly curved relatively-reciprocable cylindrical mating surfaces, the curves of which are struck from a common axis remote from the cross head, the axis of said mating cylindrical surfaces being parallel to the axis of rotation of one of said rotors.

4. In a pump or motor, a primary rotor having a cylinder, a piston reciprocable therein, a secondary roto'r having a guideway, 'a two-part piston crosshead located within said guideway and having mating cylindrical surfaces of the same curvature whose axis is parallel to the axis of rotation of one of said rotors, with the centers of curvature of said surfaces lying on the rotational axis of said primary rotor, one of said cross head parts being rigidly attached to saidv piston.

5. In a pump or motor, a primary rotor having 40 a cylinder, a .piston reciprocable therein, a secondary rotor, a crosshead guide block secured to said secondary rotor and having a. guideway therein, and a two-part -piston crosshead Klocated within said guideway and having mating. cylindrically-curved surfaces providing free sliding contacts between said crosshead and said guideway, said crosshead guide block guideway being arranged for reciprocating movement of the twopart piston crosshead relatively thereto, the axis of said mating cylindrical surfaces being parallel to theaxis of rotation of one of said rotors, one of said crosshead parts being rigidly attached to said piston.

ELEK K. BENEDEK. 

